Non-contacting face seal assemblies are usually applied to high-speed, high-pressure rotating equipment where the use of ordinary mechanical face seal assemblies with face contact would result in excessive generation of heat and wear. Non-contacting operation avoids this undesirable face contact when the shaft is rotating above a certain minimum speed, which is often called a lift-off speed.
As with ordinary contacting-type mechanical seal assemblies, a non-contacting face seal assembly consists of two sealing rings, each of which is provided with a very precisely finished sealing surface or face. These surfaces are perpendicular to and concentric with the axis of rotation. Both rings are positioned adjacent to each other with the sealing surfaces in contact at conditions of zero pressure differential and zero speed of rotation. One of the rings is normally fixed to the rotatable shaft, and the other is located within the seal housing structure and allowed to move axially. To enable axial movement of this sealing ring and yet prevent leakage of the sealed fluid, a sealing element is placed between this ring and the housing. This sealing element must permit some sliding motion while under pressure, and therefore a top quality O-ring is normally selected for that duty. This O-ring is often called the secondary seal.
To achieve non-contacting operation of the seal assembly, one of the two sealing surfaces in contact is usually provided with shallow surface recesses, which act to generate pressure fields that force the two sealing surfaces apart. When the magnitude of the forces resulting from these pressure fields is large enough to overcome the forces that urge the seal faces closed, the sealing surfaces will separate and form a clearance, resulting in non-contacting operation. The character of the separation forces is such that their magnitude decreases with the increase of face separation. Opposing or closing forces, on the other hand, depend on sealed pressure level and as such are independent of face separation. They result from the sealed pressure and the spring force acting on the back surface of the axially movable sealing ring. Since the separation or opening force depends on the separation distance between sealing surfaces, during the operation of the seal or on imposition of sufficient pressure differential equilibrium separation between both surfaces will establish itself. This occurs when closing and opening forces are in equilibrium and equal to each other. Equilibrium separation constantly changes within the range of gaps. The goal is to have the low limit of this range above zero. Another goal is to make this range as, narrow as possible, because on its high end the separation between the faces will lead to increased seal leakage. Since non-contacting seals operate by definition with a clearance between sealing surfaces, their leakage will be higher than that of a contacting seal of similar geometry. Yet, the absence of contact will mean zero wear on the sealing surfaces and therefore a relatively low amount of heat generated between them. It is this low generated heat and lack of wear that enables the application of non-contacting seal assemblies (commonly referred to as dry gas seals) to high-speed turbomachinery, where the sealed fluid is gas. Turbocompressors are used to compress this fluid and since gas has a relatively low mass, they normally operate at very high speeds and with a number of compression stages in series.
During a typical period of operation, a turbocompressor is started and the power unit starts the shaft rotating. At the initial warm-up stage of operation, shaft speeds may be quite low. Typically, oil is used to support the shaft at its two radial bearings and one thrust bearing. Oil warms up in oil pumps and also accepts shear heat from compressor bearings. The oil together with process fluid turbulence and compression in turn warm-up the compressor. Once the full operating speed is reached, the compressor reaches in time some elevated equilibrium temperature. On shutdown, shaft rotation stops and the compressor begins to cool down. In this situation, various components of the compressor cool down at different rates and, importantly, the shaft contracts with decreasing temperature at a different rate than the compressor casing. The net result of this at the seal assembly is the axial creeping motion of the shaft and the seal parts fixed to it, which may move the rotatable sealing face away from the stationary sealing face. With often only a spring load behind the stationary sealing ring, the stationary sealing face may not be able to follow the retracting rotating face, if the above-mentioned secondary seal has too much friction. These prior art secondary seal arrangements can be found for example in U.S. Pat. Nos. 4,768,790; 5,058,905 or 5,071,141. The term used often in the industry for this phenomenon is "seal face hang-up". In such case there may be a very high leakage of process fluid the next time the compressor is restarted and often in such cases the seal assembly will resist all attempts to reseal it. The seal assembly must then be removed and replaced at a considerable cost in time and lost production.
This invention is aimed at the reduction of friction forces at the secondary seal of a non-contacting face seal assembly to prevent its excessive drag and the corresponding hang-up of the axially movable seal face which causes high process fluid leakage. These friction forces cannot be lowered beyond a certain value with prior art arrangements, where typically an O-ring or a similar elastomer seal would be placed between two concentric cylindrical surfaces. While these surfaces can be machined with a high degree of accuracy to provide for uniform radial clearance to accept the secondary seal, the elastomer-type secondary seal itself is typically quite non-uniform in its cross-section. To eliminate the possibility of leakage, it is then necessary to design the radial clearance for this secondary seal narrower than what is the cross-sectional dimension of the secondary seal at its thinnest point. Given the relatively high non-uniformity of O-ring or similar seal cross-sections, this results in considerable squeeze in areas where the secondary seal is thicker, and therefore results in considerable friction and drag which can cause hang-up.
Another aim of the invention is to assure a reliable sealing contact despite secondary seal cross-section non-uniformities.
The improvement this invention provides is a compliant spring element cooperating with the O-ring or similar secondary seal. The prior art requirement to squeeze the secondary seal into a uniform radial gap with the consequence of high friction forces is thus eliminated. Circumferential compliance of the spring enables the spring to place a considerably lower and more uniform load onto the secondary seal, a load which is relatively independent of variations in the secondary seal cross-section. This results in dramatically lower friction and drag forces and therefore a lesser danger of the seal face hang-up, resulting in a more reliable sealing action.
Other objects and purposes of the invention will be apparent from the detailed description of the invention as presented below.